US5715865A - Pressure compensating hydraulic control valve system - Google Patents

Pressure compensating hydraulic control valve system Download PDF

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Publication number
US5715865A
US5715865A US08/747,843 US74784396A US5715865A US 5715865 A US5715865 A US 5715865A US 74784396 A US74784396 A US 74784396A US 5715865 A US5715865 A US 5715865A
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United States
Prior art keywords
pressure
chamber
valve
poppet
hydraulic
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Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Expired - Fee Related
Application number
US08/747,843
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English (en)
Inventor
Raud A. Wilke
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Husco International Inc
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Husco International Inc
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Application filed by Husco International Inc filed Critical Husco International Inc
Priority to US08/747,843 priority Critical patent/US5715865A/en
Assigned to HUSCO INTERNATIONAL, INC. reassignment HUSCO INTERNATIONAL, INC. ASSIGNMENT OF ASSIGNORS INTEREST (SEE DOCUMENT FOR DETAILS). Assignors: WILKE, RAUD A.
Priority to EP97909916A priority patent/EP0877975B1/en
Priority to JP52255398A priority patent/JPH11502048A/ja
Priority to BR9707130A priority patent/BR9707130A/pt
Priority to KR1019980705345A priority patent/KR100298068B1/ko
Priority to DE69705498T priority patent/DE69705498T2/de
Priority to PCT/US1997/017691 priority patent/WO1998021632A1/en
Priority to CA 2241248 priority patent/CA2241248A1/en
Publication of US5715865A publication Critical patent/US5715865A/en
Application granted granted Critical
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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B13/00—Details of servomotor systems ; Valves for servomotor systems
    • F15B13/02—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors
    • F15B13/04—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors for use with a single servomotor
    • F15B13/0416—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors for use with a single servomotor with means or adapted for load sensing
    • F15B13/0417—Load sensing elements; Internal fluid connections therefor; Anti-saturation or pressure-compensation valves
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B13/00—Details of servomotor systems ; Valves for servomotor systems
    • F15B13/02—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors
    • F15B13/04—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors for use with a single servomotor
    • F15B13/0401—Valve members; Fluid interconnections therefor
    • F15B13/0405—Valve members; Fluid interconnections therefor for seat valves, i.e. poppet valves
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F16—ENGINEERING ELEMENTS AND UNITS; GENERAL MEASURES FOR PRODUCING AND MAINTAINING EFFECTIVE FUNCTIONING OF MACHINES OR INSTALLATIONS; THERMAL INSULATION IN GENERAL
    • F16K—VALVES; TAPS; COCKS; ACTUATING-FLOATS; DEVICES FOR VENTING OR AERATING
    • F16K17/00—Safety valves; Equalising valves, e.g. pressure relief valves
    • F16K17/20—Excess-flow valves
    • F16K17/22—Excess-flow valves actuated by the difference of pressure between two places in the flow line
    • G—PHYSICS
    • G05—CONTROLLING; REGULATING
    • G05D—SYSTEMS FOR CONTROLLING OR REGULATING NON-ELECTRIC VARIABLES
    • G05D16/00—Control of fluid pressure
    • G05D16/04—Control of fluid pressure without auxiliary power
    • G05D16/10—Control of fluid pressure without auxiliary power the sensing element being a piston or plunger
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/20—Fluid pressure source, e.g. accumulator or variable axial piston pump
    • F15B2211/265—Control of multiple pressure sources
    • F15B2211/2656—Control of multiple pressure sources by control of the pumps
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B2211/00—Circuits for servomotor systems
    • F15B2211/30—Directional control
    • F15B2211/305—Directional control characterised by the type of valves
    • F15B2211/30525—Directional control valves, e.g. 4/3-directional control valve
    • F15B2211/3053—In combination with a pressure compensating valve
    • F15B2211/30555—Inlet and outlet of the pressure compensating valve being connected to the directional control valve
    • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y10—TECHNICAL SUBJECTS COVERED BY FORMER USPC
    • Y10T—TECHNICAL SUBJECTS COVERED BY FORMER US CLASSIFICATION
    • Y10T137/00—Fluid handling
    • Y10T137/8593—Systems
    • Y10T137/87169—Supply and exhaust

Definitions

  • the present invention relates to valve assemblies which control hydraulically powered machinery; and more particularly to pressure compensated valves wherein a fixed differential pressure is to be maintained in order to maintain a uniform flow rate.
  • the speed of a hydraulically driven working member on a machine depends upon the cross-sectional area of principal narrowed orifices of the hydraulic system and the pressure drop across those orifices.
  • pressure compensating hydraulic control systems have been designed to eliminate the pressure drop. These previous control systems include sense lines which transmit the pressure at the valve workports to the input of a variable displacement hydraulic pump supplying pressurized hydraulic fluid in the system. The resulting self-adjustment of the pump output provides an approximately constant pressure drop across a control orifice whose cross-sectional area can be controlled by the machine operator. This facilitates control because, with the pressure drop held constant, the speed of movement of the working member is determined only by the cross-sectional area of the orifice.
  • One such system is disclosed in U.S. Pat. No. 4,693,272 entitled "Post Pressure Compensated Unitary Hydraulic Valve", the disclosure of which is incorporated herein by reference.
  • the "bottoming out" of a piston driving a load could cause the entire system to "hang up". This could occur in such systems which used the greatest of the workport pressures to motivate the pressure compensation system. In that case, the bottomed out load has the greatest workport pressure and the pump is unable to provide a greater pressure; thus there would no longer be a pressure drop across the control orifice.
  • such systems may include a pressure relief valve in a load sensing circuit of the hydraulic control system. In the bottomed out situation, the relief valve opens to drop the sensed pressure to the load sense relief pressure, enabling the pump to provide a pressure drop across the control orifice.
  • the present invention is directed toward satisfying those needs.
  • a hydraulic valve assembly for feeding hydraulic fluid to at least one load includes a pump of the type which produces a variable output pressure which at any time is the sum of input pressure at a pump input port and a constant margin pressure.
  • a separate valve section controlling the flow of hydraulic fluid from the pump to a hydraulic actuator connected to one of the loads and subjected to a load force that creates a load pressure.
  • the valve sections are of a type in which the greatest load pressure is sensed to provide a load sense pressure which is transmitted to the control input port of the pump.
  • Each valve section has a metering orifice through which the hydraulic fluid passes from the pump to the respective actuator.
  • a pressure compensating valve within each valve section provides the load sense pressure at the other side of the metering orifice, so that the pressure drop across the metering orifice is substantially equal to the constant amount.
  • the pressure compensator has a poppet that slides within a bore and divides the bore into first and second chambers. The first chamber communicates with the other side of the metering orifice and the second chamber is in communication with the load sense pressure. As a result changes in a pressure differential between the first and second chambers causes movement of the poppet, where the magnitude and direction of that pressure differential determines a position of the poppet with in the bore.
  • the bore has an output port from which fluid is supplied to the respective hydraulic actuator.
  • the poppet having a passage through which fluid can flow between the metering orifice and the output port with the amount of the flow governed by the position of the poppet. Such flow is enabled when pressure in the first chamber is greater than pressure in the second chamber and is disabled when the pressure in the second chamber is significantly greater than the pressure in the first chamber.
  • a check valve is located within the poppet and controls communication of pressure between the first chamber and one of the output port and the first second chamber.
  • the check valve is in the passage of the poppet and closes that passage in response to pressure at the output port being greater than pressure in the first chamber, thereby preventing back flow of fluid from the actuaor to the pump under an excessive load pressure.
  • the poppet has a pilot passage between the first and second chambers. Here the check valve closes the pilot passage in response to pressure in the second chamber being greater than pressure in the first chamber.
  • FIG. 1 a schematic diagram of a hydraulic system which incorporates a multiple valve assembly according to the present invention
  • FIG. 2 is a partially schematic, partially sectional side-view of a valve which embodies the invention.
  • FIG. 3 is an orthogonal cross-sectional view of the valve in FIG. 2;
  • FIG. 4 is a cross sectional view that is similar to FIG. 3, but of another embodiment of the present invention.
  • FIG. 1 schematically depicts a hydraulic system 10 having a multiple valve assembly 12 which controls all motions of hydraulically powered working members of a machine, such as the boom and bucket of a backhoe.
  • the valve assembly 12 comprises of several individual valve sections 13, 14 and 15 interconnected side-by-side with each section used to control one degree of movement of the working members.
  • a given valve section 13, 14 or 15 controls the flow of hydraulic fluid from a pump 16 to one of several actuators 20 connected to the working members and the return of the fluid to a reservoir or tank 18.
  • Each actuator 20 has a cylinder housing 22 within which is a piston 24 that divides the housing interior into a bottom chamber 26 and a top chamber 28.
  • the pump 16 typically is located remotely from the valve assembly 12 and is connected by a supply conduit or hose 30 to a supply passage 31 extending through the valve assembly 12.
  • the pump 16 is a variable displacement type whose output pressure is designed to be the sum of the pressure at a displacement control input port 32 plus a constant pressure, known as the "margin.”
  • the control input port 32 is connected to a transfer passage 34 that extends through the sections 13-15 of the valve assembly 12.
  • a reservoir passage 36 also extends the valve assembly 12 and is coupled to the tank 18.
  • valve sections 14 in the illustrated embodiment.
  • valve sections 13-15 in the assembly 12 operates similarly, and the following description is applicable to each one.
  • valve section 14 has a body 40 and control spool 42 which a machine operator can move in either reciprocal direction within a bore in the body by operating a control member which may be attached thereto, but which is not shown.
  • hydraulic fluid, or oil is directed to the bottom or top chamber 26 and 28 of a cylinder housing 22 and thereby drives the piston 24 up or down.
  • the extent to which the machine operator moves the control spool 42 determines the speed of the working member connected to the piston 24.
  • the machine operator moves the reciprocal control spool 42 leftward. This opens passages which allows the pump 16 (under the control of the load sensing network to be described later) to draw hydraulic fluid from the reservoir 18 and force it to flow through pump output conduit 30, into a supply passage 31 in the body 40. From the supply passage 31 the fluid passes through a metering orifice formed by spool notch 44 of the control spool 42, through feeder passage 43 and through the variable orifice 46 ( Figure 3) formed by a pressure compensating check valve 48.
  • control spool 42 To move the piston 24 downward, the machine operator moves control spool 42 rightward, which opens a corresponding set of passages so that the pump 16 forces hydraulic fluid into the top chamber 28, and out of the bottom chamber 26 of the cylinder housing 22, causing the piston 24 to move downward.
  • the machine operator would have difficulty controlling the speed of the piston 24.
  • the difficulty results from the speed of piston movement being directly related to the flow rate of the hydraulic fluid, which is determined primarily by two variables--the cross sectional areas of the most restrictive orifices in the flow path and the pressure drops across those orifices.
  • the most restrictive orifice is the metering notch 44 of the control spool 42 and the operator is able to control the cross sectional area of the orifice by moving the control spool. Although this controls one variable which helps determine the flow rate, it provides less than optimum control because flow rate is also directly proportional to the square root of the total pressure drop in the system, which occurs primarily across spool notch 44.
  • adding material to the bucket of a backhoe might increase the pressure in the bottom cylinder chamber 26, which would reduce the difference between that load pressure and the pressure provided by the pump 16. Without pressure compensation, this reduction of the total pressure drop would reduce the flow rate and thereby reduce the speed of the piston 24 even if the machine operator holds metering notch 44 at a constant cross sectional area.
  • the present invention relates to a pressure compensation mechanism that is based upon the pressure compensating check valve 48 in each valve section 13-15.
  • the pressure compensating check valve 48 has a poppet 60 which sealingly slides reciprocally in a bore 62 in the valve body 40, dividing bore 62 into a first chamber 64, which is in communication with feeder passage 43, and a second chamber 66.
  • the poppet 60 is biased downward (in the illustrated orientation) by a first spring 68 located in the first chamber 64.
  • the top side 70 and a bottom side 71 of poppet 60 have equal areas.
  • the poppet 60 has a central bore 85 with lateral apertures 87 which together form a path through the pressure compensating check valve 48 which is the variable orifice 46 referred to above.
  • the poppet 60 has an internal check valve within the central bore 85.
  • the check valve comprises a valve member 82 biased by a second spring 84 into a closed state abutting an aperture ring 86.
  • the aperture ring is held against a shoulder of the poppet bore by a snap ring 88 received within an annular groove in the bore.
  • the pressure compensating mechanism senses the pressure at each powered workport of every valve section 13-15 in the multiple valve assembly 12, selects the greatest of these workport pressures to be applied to the displacement control input port 32 of the hydraulic pump 16. This selection is performed by a chain of shuttle valves 72, each of which is in a different valve section 13 and 14.
  • the first valve section 15 in the chain need not have a shuttle valve (see FIG. 1).
  • the inputs to its shuttle valve 72 are (a) the feeder passage 43 (via shuttle passage 74) and (b) the through passage 76 of the upstream valve section 15 which has the powered workport pressures in the valves upstream from middle valve section 14.
  • the feeder passage 43 sees the pressure of the powered one of workport 54 or 56, or the pressure of reservoir passage 36 when the spool 42 is in neutral.
  • the shuttle valve 72 operates to transmit the greater of the pressures at inputs (a) and (b) via its section's through passage 76 to the shuttle valve 72 of the adjacent downstream valve section 13.
  • the through-passage 76 of the farthest downstream valve section 13 in the chain of shuttle valves 72 opens into the transfer passage 34 which is connected to the pump control input port 32. Therefore, in the manner just described, the greatest of all the powered workport pressures in the valve assembly is transmitted to the control input port 32. The greatest of the powered workport pressures also is applied via the transfer passage 34 through each valve section 13-15 to second chamber 66 of pressure compensating check valves 48, thereby exerting that pressure on the bottom 71 of poppet 60.
  • An end section 78 of the valve assembly 12 contains ports for connecting the supply passage 31, transfer passage 34 and reservoir passage 36 to the pump 16 and the tank 18. This end section also includes a pressure relief valve 80 that relieves excessive pressure in the pump control transfer passage 34 to the tank 18.
  • variable orifice path through the pressure compensating check valve 48 must be at least partially open. For this to occur, the poppet 60 must be moved downward so that lateral apertures 87 communicate with the bridge passage 50. Because the areas of bottom 71 and top 70 sides of the poppet 60 are equal, fluid flow is throttled at orifice 46 so that the pressure in the first chamber 64 of compensation valve 48 is approximately equal to the greatest workport pressure in the second chamber 66. This pressure is communicated to one side of spool metering notch 44 via feeder passage 43 in FIG. 2. The other side of metering notch 44 is in communication with supply passage 31, which receives the pump output pressure that is equal to the greatest workport pressure plus the margin.
  • the pressure drop across the metering notch 44 is equal to the margin.
  • Changes in the greatest workport pressure are seen both at the supply side (passage 31) of metering notch 44 and at the bottom side 71 of pressure compensating poppet 60.
  • the pressure compensating poppet 60 finds a balanced position so that the load sense margin is maintained across metering notch 44.
  • the operation of the pressure compensating check valve 48 causes the pump margin pressure to be the approximately constant pressure drop across the metering notch 44.
  • FIG. 4 depicts another embodiment for achieving this result without employing a shuttle valve chain.
  • a valve section 100 has the valve body 102 with a control spool (not shown) which operates in the same manner as described with respect to the previous embodiment with the feeder passage 43 from the control spool communicating with the first chamber 110 of the bore 104 of a pressure compensating check valve 106.
  • the second chamber 112 of the valve bore 104 in turn communicates with the transfer passage 34 that leads to the control input port 32 of hydraulic pump 16.
  • the pressure compensating check valve 106 includes poppet 108 which sealingly slides reciprocally in the bore 104 and divides the bore into the first and second chambers 110 and 112.
  • the top side and bottom sides of poppet 108 have equal areas.
  • the poppet 108 is biased downward (in the illustrated orientation) by a first spring 114 located in the first chamber 110.
  • a path through a central poppet bore 118 is opened between first chamber 110 and a bridge passage 116, similar to bridge passage 50 in the first embodiment. This path is the variable orifice of the valve section as described previously.
  • a pilot passage 120 extends through the poppet 108 from the bottom surface to the internal bore 118 and a check valve 122 is formed in the pilot passage.
  • the orientation of the check valve 122 is such that when the pressure in the internal bore 118 is the largest workport pressure of all the valve sections 13-15, the check valve 122 opens to apply that pressure to the transfer passage 34 and thus to the control input port 32 of the pump 16.
  • the check valve 122 closes, as shown in FIG. 4, when the workport pressure of this valve section 14 is not the greatest workport pressure in the entire multiple valve assembly 12. This occurs when the pressure in the second bore chamber 112, received via transfer passage 34 from another valve section 13 or 15, is greater than the workport pressure in poppet bore 118 of this valve section 114.

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  • Engineering & Computer Science (AREA)
  • Physics & Mathematics (AREA)
  • General Engineering & Computer Science (AREA)
  • Fluid Mechanics (AREA)
  • Mechanical Engineering (AREA)
  • General Physics & Mathematics (AREA)
  • Automation & Control Theory (AREA)
  • Fluid-Pressure Circuits (AREA)
US08/747,843 1996-11-13 1996-11-13 Pressure compensating hydraulic control valve system Expired - Fee Related US5715865A (en)

Priority Applications (8)

Application Number Priority Date Filing Date Title
US08/747,843 US5715865A (en) 1996-11-13 1996-11-13 Pressure compensating hydraulic control valve system
CA 2241248 CA2241248A1 (en) 1996-11-13 1997-09-30 Pressure compensating hydraulic control valve system
BR9707130A BR9707130A (pt) 1996-11-13 1997-09-30 Sistema hidráulico e mecanismo de válvula hidráulica
JP52255398A JPH11502048A (ja) 1996-11-13 1997-09-30 圧力補償液圧制御装置
EP97909916A EP0877975B1 (en) 1996-11-13 1997-09-30 Pressure compensating hydraulic control valve system
KR1019980705345A KR100298068B1 (ko) 1996-11-13 1997-09-30 압력보상유압제어밸브장치
DE69705498T DE69705498T2 (de) 1996-11-13 1997-09-30 Ventilsystem für druckkompensierende hydraulische steuerung
PCT/US1997/017691 WO1998021632A1 (en) 1996-11-13 1997-09-30 Pressure compensating hydraulic control valve system

Applications Claiming Priority (1)

Application Number Priority Date Filing Date Title
US08/747,843 US5715865A (en) 1996-11-13 1996-11-13 Pressure compensating hydraulic control valve system

Publications (1)

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US5715865A true US5715865A (en) 1998-02-10

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Application Number Title Priority Date Filing Date
US08/747,843 Expired - Fee Related US5715865A (en) 1996-11-13 1996-11-13 Pressure compensating hydraulic control valve system

Country Status (8)

Country Link
US (1) US5715865A (ja)
EP (1) EP0877975B1 (ja)
JP (1) JPH11502048A (ja)
KR (1) KR100298068B1 (ja)
BR (1) BR9707130A (ja)
CA (1) CA2241248A1 (ja)
DE (1) DE69705498T2 (ja)
WO (1) WO1998021632A1 (ja)

Cited By (32)

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US6098403A (en) * 1999-03-17 2000-08-08 Husco International, Inc. Hydraulic control valve system with pressure compensator
US6318079B1 (en) 2000-08-08 2001-11-20 Husco International, Inc. Hydraulic control valve system with pressure compensated flow control
US6637461B2 (en) 2002-03-08 2003-10-28 Husco International, Inc. Electrically operated hydraulic actuator with force feedback position sensing
US20030205128A1 (en) * 2002-05-02 2003-11-06 Sauer-Danfoss (Nordborg) A/S Hydraulic valve system
US20030205278A1 (en) * 2002-05-02 2003-11-06 Sauer-Danfoss (Nordborg) A/S Hydraulic valve system
US20030205279A1 (en) * 2002-05-02 2003-11-06 Sauer-Danfoss (Nordborg) A/S Hydraulic valve system
US20040216599A1 (en) * 2003-05-02 2004-11-04 Pieper Gary J. Apparatus and method for providing reduced hydraulic flow to a plurality of actuatable devices in a pressure compensated hydraulic system
US20050211320A1 (en) * 2004-03-26 2005-09-29 Jason Greenwood Hydraulic valve section with reduced bore distortion
EP1610002A1 (en) 2004-06-24 2005-12-28 Walvoil S.p.A. Saturation-proof hydraulic control device with two or more elements
US20060037649A1 (en) * 2004-08-17 2006-02-23 Walvoil S.P.A. Anti-saturation directional control valve composed of two or more sections with pressure selector compensators
US20070056437A1 (en) * 2003-08-08 2007-03-15 Volker Bosebeck Hydraulic control system for construction vehicle, particularly excavators
US20070130935A1 (en) * 2005-12-12 2007-06-14 Husco International, Inc. Integrated valve assembly and computer controller for a distributed hydraulic control system
US20070144588A1 (en) * 2005-12-23 2007-06-28 Husco International, Inc. Spool activated lock-out valve for a hydraulic actuator load check valve
US20080072749A1 (en) * 2006-09-27 2008-03-27 Pfaff Joseph L Hydraulic valve assembly with a pressure compensated directional spool valve and a regeneration shunt valve
US20090173067A1 (en) * 2008-01-09 2009-07-09 Pack Andreas S Hydraulic control valve system with isolated pressure compensation
US20100307606A1 (en) * 2009-06-09 2010-12-09 Russell Lynn A Control valve assembly with a workport pressure regulating device
US20100307621A1 (en) * 2007-11-14 2010-12-09 Rueb Winfried Hydraulic valve device
ITMO20090220A1 (it) * 2009-09-03 2011-03-04 Brevini Fluid Power S P A Valvola di distribuzione
EP2439416A2 (en) 2010-10-08 2012-04-11 Husco International, Inc. Flow summation system for controlling a variable displacement hydraulic pump
CN102536942A (zh) * 2012-02-13 2012-07-04 中联重科股份有限公司 节流回路
CN102913438A (zh) * 2011-08-05 2013-02-06 江苏国瑞液压机械有限公司 一种液压泵转换阀
WO2013096300A2 (en) 2011-12-22 2013-06-27 Husco International, Inc. Hydraulic system with fluid flow summation control of a variable displacement pump and priority allocation of fluid flow
US9145905B2 (en) 2013-03-15 2015-09-29 Oshkosh Corporation Independent load sensing for a vehicle hydraulic system
EP3246578A1 (en) 2016-05-13 2017-11-22 Robert Bosch GmbH Hydraulic system for controlling an implement
WO2018200700A1 (en) * 2017-04-28 2018-11-01 Eaton Intelligent Power Limited System for damping mass-induced vibration in machines having hydraulically controlled booms or elongate members
US10619750B2 (en) * 2014-06-25 2020-04-14 Parker-Hannifin Corporation Reverse flow check valve in hydraulic valve with series circuit
US11035389B2 (en) 2017-04-28 2021-06-15 Eaton Intelligent Power Limited Drift compensation system for drift related to damping of mass-induced vibration in machines
US11209027B2 (en) 2014-07-15 2021-12-28 Eaton Intelligent Power Limited Methods and apparatus to enable boom bounce reduction and prevent un-commanded motion in hydraulic systems
US11209028B2 (en) 2017-04-28 2021-12-28 Eaton Intelligent Power Limited System with motion sensors for damping mass-induced vibration in machines
US11326627B2 (en) 2013-08-30 2022-05-10 Danfoss Power Solutions Ii Technology A/S Control method and system for using a pair of independent hydraulic metering valves to reduce boom oscillations
US11566642B2 (en) 2013-11-14 2023-01-31 Danfoss Power Solutions Ii Technology A/S Pilot control mechanism for boom bounce reduction
US11773880B2 (en) * 2017-05-25 2023-10-03 Faster S.R.L. Connector for hydrodynamic applications equipped with at least one detection sensor

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CN105840574B (zh) * 2015-01-16 2018-04-06 徐工集团工程机械股份有限公司 负载敏感多路阀及工程机械液压系统
CN107339277B (zh) * 2017-08-31 2024-05-14 潍柴雷沃智慧农业科技股份有限公司 一种拖拉机负载敏感式多路阀

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KR100298068B1 (ko) 2001-08-07
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EP0877975A1 (en) 1998-11-18
JPH11502048A (ja) 1999-02-16
KR19990077204A (ko) 1999-10-25
DE69705498T2 (de) 2002-04-25
DE69705498D1 (de) 2001-08-09
CA2241248A1 (en) 1998-05-22
WO1998021632A1 (en) 1998-05-22

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